Hydraulic power transmission and braking system for vehicles

ABSTRACT

A hydraulic power transmission and braking system for vehicles in which individual hydraulic wheel motors are connected through suitable control apparatus to a variable displacement hydraulic pump which is driven by an appropriate power source. The displacement of the pump varies automatically in response to hydraulic fluid pressure and flow rate requirements imposed by the vehicle operating conditions. The power transmission control apparatus automatically switches from the four-wheel drive mode to the two-wheel drive mode as the vehicle accelerates past a predetermined speed. The braking control system proportions the braking force applied to each of the hydraulic wheel motors in accordance with the load imposed on that wheel by the vehicle operating conditions.

United States Patent 11 1 1111 3,827,239

Ulrich, Jr. 1451 Aug. 6, 1974 [54] HYDRAULIC POWER TRANSMISSION AND2,774,436 12/1956 Ferris 60/19 BRAKING SYSTEM FOR VEHICLES 2,941,3656/1960 Carlson et a] 60/19 [75] Inventor: Bernhard Ulrich, Jr., CorpusPrimary Examiner Edgar W Geoghegan Chm, Attorney, Agent, or Firm--Darby& Darby [73] Assignee: Tex-Trans, Inc., Corpus Christi,

Tex. [57] ABSTRACT 22 F1 d: N 10 1972 A hydraulic power transmission andbraking system 1 M for vehicles in which individual hydraulic wheel mo-PP 4 305,513 tors are connected through suitable control apparatus to avariable displacement hydraulic pump which is [52 US. (:1 60/420, 60/19,60/427, drive by POwer z displace" 60/431, 60/434, 60/448, 60/449,60/451, 60/452 ment of the pump vanes automatically 1n response to 511111. c1. Fl6h 39/46, Flb 11/16 hydraulic fluid Press"re andrequirements [58] Field of Search /19, 420 427, 431, 434, imposed by thevehicle operating conditions. The

60/44] 445 448 449 451452 power transmission control apparatusautomatically 180/66, 44 switches from the four-wheel drive mode to thetwowheel drive mode as the vehicle accelerates past a l 56] ReferencesCited predetermined speed. The braking control system proportions thebraking force applied to each of the hy- UNITED STATES PATENTS draulicwheel motors in accordance with the load im- Kenyon 6t 8] X posed onwheel the vehicle operating condi- 2,403,519 7/1946 Gardiner 60/393tions 2,416,80l 3/1947 Robinson 60/452 2,516,662 7/1950 Vickers et al180/66 R 20 Claims, 19 Drawing Figures W J 1 DIRECTION SENSOR L Q; HIGHI 62 PRESSURE 00 a 1 72 51 5a 33 A 1 i 1 75 27 23 22 i FLOW s I RATEomveu I 1 INDICATOR MAIN L VARIABLE I I I eta M501. 20 mspi c z gmau'r IQQ 3 L (PRIME MOVER) i I 16 24 A 44 k 1 D L S l 14 I 1 48 73 9 52 5 l 86i 42 FWD REV| D 1 RESERVOIR FILTER I 1 I ss l 0 9 M1 1 2,, 1 PRESSURE LI I REGULATOR ---+l I 1 1 I L e24 i l L l. 'E

P51) L SCAVENG E (NEAR ATMOSPHERIC PRESSURE) PATENTED AUG 6 1974 SHEU 0BOF H PAIENTEDAUE 6l974 SHEET m m w $827239 Pmmwws- 61914 3.827. 239

srmmnnp TO RIGHT TO RIGHT FRONT BRAKE F I REAR BRAKE FWD REV,

FWD QEV FRONT OF VEHICLE REV FWD REV FWD TO LEFT 452 TO LEFT FRONT BRAKEREAR BRAKE HYDRAULIC POWER TRANSMISSION AND BRAKING SYSTEM FOR VEHICLESThis invention relates to hydraulic power transmis sion systems and,more particularly, to a hydraulic power transmission and braking systemfor vehicles such as passenger automobiles.

One of the problems of the prior art power transmission systems forvehicles is that the prime mover, or engine, of the vehicle mustoperateover a relatively wide range of speeds in order to accommodate the widerange of conditions under which a typical vehicle, such as, for example,a passenger automobile, is operated. For example, during theacceleration of an automobile equipped with a conventional manual orautomatic transmission system, the speed of the automobiles enginevaries over a relatively wide speed range and only momentarily operatesat the speed at which it produces its maximum output power level.

Another problem of conventional vehicle power transmission systems isthat they are relatively heavy and bulky, their propeller shafts anddifferential present problems of road clearance and, in the case ofpassenger automobiles, a tunnel through the passenger compartment isoften required. Further, conventional automatic transmission systems arecomplex and expensive to build and maintain.

Further, conventional vehicular power transmission systems are normallyboth structurally and functionally independent from the braking systemsused in the vehicles, with the result that, if either the powertransmission system or the braking system fails, the vehicle isdisabled.

Moreover, conventional braking systems fail to apportion the brakingforce to each wheel in accordance with the load on that wheel. Forexample, when automobile brakes are applied sharply, the momentum of thevehicle causes an increased load on the front wheels, and a reduced loadon the rear wheels, but conventional braking systems apply approximatelyequal braking force to both the front wheels and the rear wheels, withthe result that the lightly loaded rear wheels tend to lock and skid.Similarly, conventional braking systems fail to compensate for wheelloading conditions that occur when the vehicle is operated on an inclineor rounding a corner, or when the vehicle is unbalanced by a heavycargo.

It is therefore an object of this invention to provide a vehicular powertransmission and braking system which overcomes the problems of theprior art systems.

It is a more particular object of this invention to provide a hydraulicpower transmission system which enables the prime mover of a vehicle tooperate at maximum horsepower at full throttle or maximum fuelefficiency at partial throttle.

It is also an object of this invention to provide a hydraulic powertransmission system using a variable displacement pump which respondsautomatically to vehicle operating conditions.

It is another object of this invention to provide a hydraulic powertransmission system that switches automatically from the four-wheeldrive mode to the twowheel drive mode when the vehicle accelerates pasta predetermined speed.

It is still another object of this invention to provide a combinedhydraulic power transmission and braking system for vehicles.

It is yet another object of this invention to provide a hydraulic powertransmission and braking system in which the braking force isautomatically apportioned between the vehicle wheels in accordance withthe load on each wheel.

According to the above and other objects, the present invention providesa hydraulic power transmission and braking system including a pluralityof individual hydraulic wheel motors, preferably one motor for eachvehicle wheel, and a variable displacement hydraulic pump which isdriven by suitable prime mover, such as, for example, an internalcombustion engine, and which is connected to the hydraulic wheel motorsthrough suitable control apparatus including a main control valve forcontrolling the flow of hydraulic fluid in response to aforward/neutral/reverse control and an accelerator control, a device forswitching the mode of operation of the system from four-wheel drive totwowheel drive when the vehicle accelerates past a predetermined speedand a braking control system which operates by restricting the flow ofhydraulic fluid from the hydraulic wheel motors, which act as pumps whenthe vehicle is decelerating.

Other objects and advantages of the present invention will be apparentfrom the following detailed description and accompanying drawings whichset forth by way of example, the principles. of the present inventionand the preferred embodiment for carrying out those principles.

IN THE DRAWINGS:

FIG. 1, shown in two parts 1A and 1B, is a schematic diagram of apreferred embodiment of the hydraulic power transmission and brakingsystem of the present invention;

FIG. 2A is a diagram of a preferred embodiment of the control apparatusof the present hydraulic power transmission system shown in the NEUTRALcondition;

FIG. 2B is a diagram of the control apparatus of FIG. 2A, shown in theFORWARD condition;

FIG. 2C is a diagram of the control apparatus of FIG. 2A, shown in theREVERSE condition;

. FIG. 3 is a cross sectional view of a preferred form of the hydraulicwheel motor used in the hydraulic power transmission and braking systemof the present invention;

FIG. 4 is a cross-sectional view of the hydraulic wheel motor takenalong the line 44 of FIG. 3;

FIG. 5 is a cross-sectional view of the exhaust and intake manifolds ofthe hydraulic wheel motor taken along the line 5-5 of FIG. 3;

FIG. 6 is a cross-sectional view of a preferred embodiment of thevariable displacement hydraulic pump used in the hydraulic powertransmission and braking system of the present invention;

FIG. 7 is a cross-sectional view of the variable displacement hydraulicpump taken along the line 7-7 of FIG. 6;

FIG. 8 is a cross-sectional view of the variable displacement pump takenalong the line 8-8 of FIG. 7 showing the coupling device used to preventrotation of the orbiting element;

FIG. 9 is a crosssectional view of the variable displacement hydraulicpump taken along the line 9--9 of FIG. 7 showing the inner and outereccentric carriers which control the displacement of the hydraulic pump;

FIG. is a cross-sectional view of the variable displacement hydraulicpump taken along the line 10-10 of FIG. 7 showing the automatic controlmechanism for the inner and outer eccentric carriers;

FIG. 11 is a cross-sectional view of the variable displacement hydraulicpump taken along the line 1111 of FIG. 7 showing the damper device forthe automatic displacement control mechanism shown in FIG. 10;

FIG. 12 is a cross-sectional view taken along the line 12-12 of FIG. 10;

FIG. 13 is a cross-sectional plan view, partly broken away, of the brakecontrol apparatus of the hydraulic power'transmission and braking systemof the present invention;

FIG. 14 is a cross-sectional view of the brake control apparatus takenalong the line 14- 14 of FIG. 13',

FIG. 15 is a cross-sectional view of the preferred form of brake valveused in the hydraulic power transmission and braking system of thepresent invention; and

FIG. 16 is a cross-sectional view of a fluid flow direction sensingdevice suitable for use in the hydraulic power transmission and brakingsystem of the present invention.

Referring in detail to FIGS. 1A and 1B of the drawings, there is shown aschematic diagram of the preferred embodiment of the hydraulic powertransmission and braking system of the present invention. In thepreferred embodiment shown in FIG. 1A, hydraulic fluid under pressure issupplied by variable displacement hydraulic pump 21 which is driven by asuitable prime mover such as, for example, an internal combustionengine. The preferred form of the variable displacement hydraulic pump21 is described in detail in connection with FIGS. 6-12. It will beappreciated, however, that other forms of variable displacementhydraulic pumps can be employed within the spirit and scope of thepresent invention. Further, it will be appreciated that certain objectsand advantages of the present invention can, in fact, be accomplished byusing a fixed displacement hydraulic pump as the source of hydraulicfluid under pressure for the system.

Briefly, the use of the variable displacement hydraulic pump 21 in thepresent hydraulic power transmission system enables more efficient useof the engine which is used as the prime mover for the vehicle. Moreparticularly, as the vehicle accelerates from a standstill, with fullthrottle opening, its engine is able to operate at the constant speed atwhich it produces its maximum power output, while the variabledisplacement pump 21 accomplishes the necessary changes in the speed andtorque ratios between the engine and the wheels.

For example, when the vehicle is starting from a standstill with fullthrottle, the variable displacement pump 21 will operate at its minimumdisplacement and maximum pressure (2,000 p.s.i., for example) so as todeliver full power to the wheels in the form of maximum torque atminimum speed. As the speed of the vehicle increases, the displacementof variable displacement pump 21 will increase and its output pressurewill drop so that a constant power level, corresponding to the maximumpower output of the engine will be delivered to the wheels as thevehicle accelerates.

In the preferred embodiment of the present invention, the ratio of themaximum displacement of variable displacement pump 21 to its minimumdisplacement is about 2.511 while the ratio of the maximum outputpressure at minimum displacement to the maximum output pressure atmaximum displacement is similarly about 2521. More specifically, themaximum output pressure at minimum displacement may be in theneighborhood of 2,000 p.s.i. while the maximum output pressure atmaximum displacement may be in the neighborhood of about 800 p.s.i. Itwill be appreciated, however, that the range of maximum to minimumdisplacement of the variable displacement pump 21 may be greater than orless than 2.5:1 depending upon the particular vehicle performance goalsto be accomplished.

Similarly, while the maximum displacement of the pump 21 may beapproximately equal to the displacement of two wheel motors in apassenger vehicle transmission system, it will be appreciated that thedisplacement of the pump 21 may be much smaller in relation to thedisplacement of the wheel motors in the case of large earth movingequipment, for example, and that the displacement of the pump may bemuch larger in relation to the motors in certain other applications suchas machine tools, for example.

The high pressure hydraulic fluid from variable displacement pump 21 isfed via line 20 to the main control valve 22 which is described ingreater detail in connection with FIGS. 2A-C of the drawings. Briefly,main control valve 22 responds to a selector control lever 101 (FIGS.2A-C) to direct the high pressure hydraulic fluid from pump 21 to line23 if the selector control lever 101 is in FORWARD position, and to line24 if the selector control lever 101 is in REVERSE position. The maincontrol valve 22 also responds to the vehicle accelerator control pedal151 (FIGS. 2A-C) and the flow rate indicator 25 to control the flow rateof hydraulic fluid to the hydraulic wheel motors 31, 32, 36 and 37 aswill be explained in greater detail in connection with FIGS. 2A-C.

Although the hydraulic wheel motors 31, 32, 36 and 37 are primarilyreferred to as motors throughout the present specification, it will beappreciated that these devices are simply hydraulic machines which canact either as motors" or as pumps depending upon operating conditions.For example, hydraulic machines 31, 32, 36 and 37 act as motors when thevehicle is accelerating and act as pumps" when the vehicle isdecelerating.

Briefly, flow rate indicator 25, which is shown in greater detail inFIGS. 2A-C, serves to increase the flow of hydraulic fluid to the wheelmotors 31, 32, 36 and 37 as the vehicle speed increases. Also, in thepreferred embodiment of the present hydraulic power transmission andbraking system, the flow rate indicator 25 serves to actuate a switch136 which oper ates the front wheel recirculating valve 26 to change themode of operation of the vehicle from four-wheel drive to two-wheeldrive or vice versa, as will be explained in greater detail inconnection with FIGS. 2A-C.

Assuming, for purposes of illustration, that the selector control leveris in FORWARD position, hydraulic fluid under pressure passes fromcontrol valve 22 via line 23, flow rate indicator 25, and lines 27, 28and 29 to the rear wheel motors 31 and 32, and via lines 23,

33, front wheel recirculating valve 26 and lines 34 and 35 to the frontwheel motors 36 and 37. in the preferred embodiment of the presentinvention, the wheel motors 31, 32, 36 and 37 are preferably fixeddisplacement hydraulic machines of the type described in greater detailin connection with P168. 3 5 of the drawings. it will be appreciated,however, that variable displacement hydraulic wheel motors may beemployed within the spirit and scope of the present invention.

The hydraulic fluid from the rear wheel motors 31 and 32 returns vialines 41, 42, 43 and 24 through main control valve 22 and line 44 to thevariable displacement pump 21. The hydraulic fluid from front wheelmotors 36 and 37 returns via lines 46 and 47 through front wheelrecirculating valve 26 and, via lines 48 and 24, through main controlvalve 22 and line 44 to variable displacement pump 21, thus completing aclosed circuit for the hydraulic fluid.

It will be apparent that, if the selector control lever 101 (FIG. 2A-C)is in REVERSE position, the high pressure hydraulic fluid will pass frommain control valve 22 via lines 24, 43, 42 and 41 to the rear wheelmotors 32 and 31, and via lines 24, 48, 47 and 46 to the front wheelmotors 37 and 36. In this case, lines 29, 28, 27, flow rate indicatorand lines 23 and 44 serve as the return path for the hydraulic fluidfrom the rear wheel motors 31 and 32 to the variable displacement pump21, while the lines 35, 34, front wheel recirculating valve 26 and lines33, 23 and 44 provide the return path for the hydraulic fluid from frontwheel motors 36 and 37 to the variable displacement pump 21.

According to the preferred form of the present invention, braking isaccomplished by means of brake valves in the return lines between eachof the wheel motors and the variable displacement pump 21. For example,if the vehicle is moving forward, braking is accomplished by brakevalves 51 and 52 which are located in the return lines 41 and 42 fromrear wheel motors 31 and 32 respectively, and by brake valves 56 and 57located in the return lines 46 and 47 from front wheel motors 36 and 37respectively. Similarly, if the vehicle is moving backward, braking isaccomplished by brake valves 54 and 55 located in the return lines 34and from front wheel motors 36 and 37 respectively, and by brake valves58 and 59 located in return lines 28 and 29 from rear wheel motors 31and 32 respectively.

Although the preferred embodiment of the present invention includes aseparate brake valve for each wheel motor, it will be appreciated thatthe braking function could be accomplished by forward brake valves inlines 43 and 48 and reverse brake valves in lines 27 and 33 or by asingle forward brake valve in line 24 and a single reverse" brake valvein line 23.

Briefly, the desired braking effect is accomplished by the degree ofclosing the appropriate brake valves so as to create back pressure onthe wheel motors which, it will be remembered, act as pumps when thevehicle is decelerating. The appropriate selection of the forwardbraking valves 51, 52, 56 and 57 or the rear braking valves 54, 55, 58and 59 is accomplished by a flow direction sensor or indicator 61 and abraking control system shown in detail in FlGS. 13 and 14. The flowdirection sensor or indicator 61 is preferably located in one of thehydraulic lines connected to the rear wheel motors 31 and 32 and simplyserves to indicate whether the vehicle is moving in a forward orrearward direction. The operation of the braking control system brakevalves and flow direction indicator will be explained in detail inconnection with FIGS. 13 16.

Overload valves 62, 63, 64 and 65 which may be of a commerciallyavailable type, are provided between the pairs of hydraulic linesserving each of the wheel motors 31, 32, 36 and 37, in order to preventpossible damage to the system due to excessive back pressure resultingfrom unusual road conditions during braking. For example, if one wheelshould strike a bump or pothole during severe braking, a large pressureimpulse might be generated at the outlet of the wheel motor which, asexplained above, acts as a pump when the vehicle is decelerating. Inorder to prevent such a pressure impulse from damaging the system, eachof the overload valves 62, 63, 64 and 65 serves to limit the pressuredifferential between the two hydraulic lines serving its associatedwheel motor to a predetermined level such as, for example, 3,000 p.s.i.If the back pressure in one or more of the return. lines 41, 42, 46 and47 exceeds the pressure on feed lines 28, 29, 34 and 35 by more than3,000 p.s.i. one or more of the overload valves 62, 63, 64 and 65 willoperate to allow hydraulic fluid to pass from the return line 41, 42, 46and 47 to their associated feed lines 23, 29, 34 and 35 respectivelythus relieving the excessive pressure.

It will be appreciated, however, that in the preferred embodiment of thepresent invention the pressure settings of the overload valves 62, 63,64 and 65 would be sufficiently high that the overload valves would notbe operated under normal road conditions even if the braking force weresufficient to lock the wheels of the vehicle.

In addition, a plurality of check valves 71 80 are provided in order toprevent the pressure in the main hydraulic circuit from falling below apredetermined level, such as for example 150 p.s.i. Excessively lowpressure in the main hydraulic circuit might cause bubbles to form inthe hydraulic fluid as a result of cavitation or as a result of airbeing drawn into the system to one or more of the many air-fluid sealsin the system. For example, if the vehicle is moving forward at arelatively high rate of speed and the brakes are applied, the pressurein one or more of the feed lines 27, 28, 29, 33, 34 and 35 may dropsubstantially. If, for example, the pressure in line 27 drops below thepredetermined level (150 p.s.i.), fluid from low pressure line 82 willpass through check valve to line 27, thus maintaining the pressure inline 27 at a minimum of 150 p.s.i., and thereby preventing the formationof bubbles. Similarly, if the pressure in line 28 should drop below thepredetermined low pressure level, hydraulic fluid would pass from thelow pressure line 82 through the check valve 72 to line 28. The othercheck valves 71 operate in a similar manner under various conditions aswill be apparent to those skilled in the art. The check valves 71 80 maybe of a conventionall commercially available type.

The preferred embodiment of the present hydraulic power transmission andbraking system also includes a scavenging system for recovering thehydraulic fluid which unavoidably leaks from the hydraulic wheel motors31, 32, 36 and 37 and the variable displacement pump 21. The hydraulicfluid is collected via line 83 at near atmospheric pressure and fed to asuitable reservoir 84 from whence it is passed through a filter 85 to apump 86 which pumps the scavenged hydraulic fluid into the low pressureline 82. A pressure regulator 87,

which may be of conventional design, controls the pressure differential(preferably about 150 p.s.i.) between the scavenging line 83 and the lowpressure line 82. An accumulator 88 is provided in the low pressure line82 to serve as a buffer between the check valves 71 80 and the pump 86.It will be apparent to those skilled in the art that the hydraulic fluidscavenged from the high pressure circuit by the scavenging line 83 isreturned to the main high pressure circuit via the pump 86, low pressureline 82, and the check valves 71 80, thus ensuring that the highpressure circuit will be properly filled with hydraulic fluid at alltimes despite the inevi table leakage in the wheel motors 31, 32, 36 and37 and the variable displacement pump 21.

The Control Mechanism Referring to FIGS. 2A, 2B and 2C of the drawings,there is shown a diagram of a preferred embodiment of the controlmechanism of the present hydraulic power transmission and brakingsystem. More particularly, FIG. 2A shows the control mechanism in theNEU- TRAL condition, FIG. 2B shows the control mechanism in the FORWARDcondition, and FIG. 2C shows the control mechanism in the REVERSEcondition. Although FIGS. 2A-C show an embodiment of the controlmechanism in which the necessary functions are performed primarily by aparticular arrangement of mechanical links, it will be appreciated bythose skilled in the art that other arrangements of mechanical links orelectromechanical devices or electronic circuits or fluid logic devicesor the like may be employed to perform the necessary control functionswithin the spirit and scope of the present invention.

Referring in detail to FIG. 2A of the drawings, there is shown a diagramof the control mechanism in the NEUTRAL condition. The selector controllever 101 is shown in the NEUTRAL position which causes an electricalcircuit to be established through contact 102A mounted on gear lever101, stationary contact 102B, and wire 103 to energize the coil 104A ofneutral solenoid 104. The energization of solenoid coil 104A moves thesolenoid plunger 1048 to the right, as shown in FIG. 2A, so as tooperate pincers 105 about fixed pivot pin 106, thus causing the jaws105A and 1053 of pincers 105 to locate movable pin 107 at apredetermined central position.

The centralization of pin 107 serves to centralize pilot valve member108 which in turn centralizes main control valve member 109 of maincontrol valve 22. The centralized position of pilot valve member 108serves to block the low pressure (150 p.s.i.) hydraulic fluid in passage1 11 from flowing via passage 1 13 to the upper portion 114 of the mainvalve chamber or via passage 116 to the lower portion 117 of the mainvalve chamber.

The centralized position of main control valve member 109 allows thehydraulic fluid from the pump 21 via conduit 20 to circulate viapassages 122 and 123 or via passages 125 and 126 in main control valveblock 112 back to the pump 21 via conduit 44. Similarly, the hydrauliclines 23 and 24 connecting the main control valve 22 to the wheel motors31, 32, 36 and 37 are each connected simultaneously to the outlet ofpump 21 via conduit 20 and the intake of pump 21 via conduit 44. Thus,when the main control valve member 109 is in the neutral position asshown in FIG. 2A, the hydraulic wheel motors 31, 32, 36 and 37 of thevehicle are effectively decoupled from the hydraulic pump 21 by reasonof the free circulation of hydraulic fluid within the main control valveblock 112.

In the embodiment shown in FIGS. 2A-C, the flow rate indicator 25includes a tapered passage 131 which is located in the high pressurehydraulic fluid circuit between the pump 21 and the rear wheel motors 31and 32 as explained in connection wiith FIG. 1A. A movable member 132having a suitable bulge 132A is disposed for longitudinal movementwithin passage 131 and is preferably biased toward the small end ofpassage 131 by a suitable compression spring 133. The position ofmovable member 132 within passage 131 is affected by the rate of flow ofthe hydraulic fluid which acts on the bulge 132A to move the movablemember 132 toward the large end of passage 131 against the force ofcompression spring 133. Because the rate of flow of hydraulic fluidthrough passage 131 is determined by the rate of flow of hydraulic fluidthrough rear wheel motors 31 and 32, it will be appreciated that theposition of movable member 132 within passage 131 is proportional tovehicle speed.

Assuming that the vehicle is at rest so that there is no flow ofhydraulic fluid through passage 131, movable member 132 will be in itslowermost or standstill position as indicated by line 134 shown in FIG.2A. In this position, an electrical circuit is established betweenelectrical contact member 135 mounted on movable member 132 andelectrical contact member 136 mounted on link 167, thus energizing coil137A of solenoid 137. The plunger 13713 of solenoid 137 is connected tothe pilot valve member 138 of front wheel recirculating valve 26, sothat, when solenoid coil 137A is energized, pilot valve member 138 ismoved to its upper position thus allowing hydraulic fluid from the lowpressure p.s.i.) system to flow via passages 141 and 142 to the upperportion 143A of the valve chamber, thus causing the valve member 144 tomove to its lower position, as shown in FIG. 2A, thereby connectinghydraulic fluid conduits 33 and 48 from main control valve 22 toconduits 147 and 148, respectively, which are connected to the frontwheel motors 36 and 37 of the vehicle.

Referring now to FIG. 2B of the drawings, there is shown a diagram ofthe control apparatus of the present hydraulic power transmission andbraking system with the selector level 101 in the FORWARD position andthe vehicle moving at relatively high speed. When the lever 10] is movedfrom the NEUTRAL position shown in FIG. 2A to the FORWARD position shownin FIG. 2B, the electrical circuit between contacts 102A and 1023 isbroken, thus deenergizing coil 104A of solenoid 104, thereby allowingplunger 104B to be retracted by the action of tension spring 1041C, thuscausing the jaws 105A and 10513 of pincers 105 to open, thereby freeingpin 107 for movement. It will be understood, however, that, initially,when the vehicle is at a standstill and before the accelerator controlpedal 151 is depressed, there will be little or no motion of pin 107when it is released by pincers 105.

The accelerator control pedal 151 is pivotally mounted on a fixed pin152 and connected by a link 153 to the throttle 154 of the vehiclesengine which drives the hydraulic pump 12 as explained above inconnection with FIG. 1A. When the accelerator control pedal 151 isdepressed, as shown in FIG. 2B, the throttle 154 is opened thus applyingpower to the hydraulic pump 2i which is connected via conduits 2d and Mto the main control valve 22.

Accelerator pedal lldl is also connected by a pin T55 to a link T56which is connected to the link T57 by means of the pin 155i mounted onlink ll5d which engages the scissors coupling arrangement i159 mountedon link One end of link i5? is connected by a pin 161 to the valvemember i312 of flow rate sending valve while the other end of link 157is connected by links 162 and it'd to the pilot valve member TM and mainvalve member MW of main control valve 22.

Briefly, when accelerator control pedal 115i is depressed, its motion istransmitted via link 1153 to open the throttle llfid of the engine, thusapplying power to the hydraulic pump 211. At the same time, the motionof pedal 511 is transmitted through link i561 and scissors couplingarrangement 1159 to link 157, thus causing link 1157 to pivot around pinlltill. The downward motion of the opposite end of link l5? istransmitted through links 162 and M3 to the pilot valve member 10% ofmain control valve 22.

As the pilot valve member lltih moves downward, the low pressure (150psi.) line llll is connected to pas sage lie within valve block H2, thusadmitting hydraulic fluid to the lower portion ll"! of the valve chamberand thereby causing the main valve member 1109 to move upward. As thevalve member MW moves upward, the outlet conduit 2% from the hydraulicpump 21 is gradually connected via passage T22 to the hydraulic line 23which is connected to the wheel motors. At the same time, hydraulic line2d from the wheel motors is gradually connected via passage 126 to thereturn conduit dd to the hydraulic pump 211 thus establishing a circuitfor the high pressure hydraulic fluid so as to drive the vehicle in theforward direction.

It will be appreciated that, as main control valve member Th9 movesupward, its motion is coupled through link lltifi to pilot valve memberMid, thus tending to centralize the pilot valve member llllh which inturn has the effect of stabilizing the position of main control valvemember W9. As a result of this follow-up relationship between main valvemember lllil and pilot valve member lltlh, the position of main valvemember 109 may become stabilized at a position somewhere between theNEUTRAL position shown in PEG. 2A and the fully raised (or FORWARD)position shown in FIG. 28 if the accelerator control pedal 115i is lessthan fully depressed. However, if the accelerator pedal l5ll is fullydepressed, it will be found that the main control valve member W9 willmove to its fully raised (or FOIL WARD) position shown in H6. 28.

It will be appreciated that, when the vehicle is at a standstill, theflow rate of the hydraulic fluid from the hydraulic pump 23 through thewheel motors Bil, 3'2, 36 and 37 and back to the hydraulic pump 211 iszero. As the vehicle begins to move forward, hydraulic fluid begins toflow through the wheel motors, and, as the speed of the vehicleincreases, the hydraulic fluid flow rate increases in direct proportion.in the preferred form of the present invention the increasing hydraulicfluid flow rate is capable of being supplied by the automaticallyincreasing displacement of the hydraulic pump 2i while the speed of thepump 211 can remain approximately constant at a level corresponding, forexample to the speed at which the vehicles engine achieves its maximumpower output.

llh

More specifically, in the preferred embodiment of the present inventionthe variable displacement hydraulic pump 2i is coupled directly to thevehicles engine so that the speed of the pump 2i is determined by thespeed of the engine which responds substantially immediately to thethrottle which is controlled by the vehicle operator. For example, ifthe operator decides to use maximum power to accelerate from astandstill, he opens the throttle thus causing the engine and pump 211to quickly speed up to full speed while the vehicle, because of itsinertia, barely begins to move forward. Under such conditions theautomatic displacement control mechanism of the pump 2i operates toreduce the displacement of the pump so as to produce a low hydraulicfluid flow rate commensurate with the slow speed of the wheel motors andto produce the maximum output pressure so as to deliver maximum torqueto the wheels. As the vehicle accelerates, the automatic displacementcontrol mechanism of the pump 2i operates to increase the displacementof the pump 21 and correspondingly reduce its output pressure so as tocontinue to deliver maximum power to the wheels through a range ofvehicle speeds while the speed of the vehicles engine and pump 21 remainapproximately constant at the maximum power output level.

it will be appreciated that, if the operator does not wish to usemaximum power, he will only partially open the throttle resulting in asomewhat lower engine and pump speed. The displacement control mechanismof the pump 2i will automatically adjust the displacement of the pump2ll to coordinate the engine speed to the wheel speed.

Referring again to FlG. 2B of the drawings, it will be seen that theflow rate of hydraulic fluid through tapered passage 13E of flow rateindicator 25 will increase in direct proportion to the increaing forwardspeed of the vehicle. The increasing fluid flow rate causes anincreasing upward pressure on bulge 132A of member T32, thus causing themovable member M2 to move upward against the force of compression spring1133. The tapered contour of passage ll3ll allows the movable member 132to find an equilibrium position corresponding to the hydraulic fluidflow rate which in turn corresponds to the vehicles speed.

The upward movement of movable member T32 is transmitted through pin leito link il57 which pivots about pin H8 and thus causes the opposite endof link 3157 to move downward. The downward movement of the opposite endof link lid? is transmitted to links 162 and 163 to the pilot valvemember 1% which operates to further raise main control valve member W9thus increasing the flow of hydraulic fluid to the wheel motors viapassage T22 and hydraulic line 23.

As the speed of the vehicle increases through the low end of the speedrange, the valve member llfi2 continues to rise within the passage ldi.of flow rate sending valve 25, but electrical contact member T35 remainsin contact with electrical contact member 11% so that solenoid coilllIi7A remains energized so that hydraulic fluid from the pump 2icontinues to be transmitted through front wheel recirculating valve 26to the front wheel motors 3d and 3'7 of the vehicle. However, as thespeed of the vehicle increases beyond a predetermined value, such as forexample, the half-speed point, movable member B2 of flow rate sendingvalve 25 rises to a point where the contact between electrical contactmembers 1% and 11% is broken, as shown in H6. 2B,

thus deenergizing the solenoid coil 137A, thus allowing the compressionspring 165 to move the solenoid plunger 1378 and pilot valve member 138to their downward position as shown in FIG. 2B. The downward position ofpilot valve member 138 allows hydraulic fluid from the low pressure (150p.s.i.) system to flow through passages 141 and 156 to the lower portion143B of the valve chamber, thus causing the main valve member 144 offront wheel recirculating valve 26 to move to its upward position asshown in FIG. 2B.

When the valve member 144 is in its upward position, the hydraulicconduits 33 and 48, which are connected to control valve 22, are blockedby portions 144A and 14413, respectively, of valve member 144. At thesame time, the passages 147 and 148, which are connected to the frontwheel motors 36 and 37, are connected together through the centralportion 143C of the valve of chamber, thus allowing the hydraulic fluidwhich moves through the front wheel motors 36 and 37 as a result of thevehicle motion to simply recirculate through the valve 26. Thus, theoperation of front wheel recirculating valve 26 when the vehicle reachesthe halfspeed point serves to automatically convert the mode ofoperation of the present hydraulic power transmission system fromfour-wheel drive to twowheel drive.

One effect of the conversion of the mode of operation of the presenthydraulic power transmission system from four-wheel drive to two-wheeldrive is to reduce by one-half the total displacement of the hydraulicwheel motors connected to the variable displacement hydraulic pump 21.At the same time, the torque reaction on the rear wheel hydraulic motors31 and 32, which remain connected to the variable displacement pump 21,is approximately doubled because the entire force of acceleration, whichwas formerly distributed between all four wheels, must now be suppliedby the rear wheels alone. As a result, the back pressure of the rearwheel hydraulic motors 31 and 32, as seen by the variable displacementhydraulic pump, is approximately doubled.

It will be appreciated that the new conditions of pressure anddisplacement caused by the conversion from four-wheel drive to two-wheeldrive will cause the variable displacement hydraulic pump 21 toautomatically reduce its displacement by one-half and correspondinglydouble its output pressure as will be explained in greater detail inconnection with FIGS. 6-12. After conversion from four-wheel drive totwo-wheel drive, the vehicle will continue to accelerate, and the displacement of the variable displacement hydraulic pump 21 will continueto increase until the vehicle reaches full speed, at which point themovable member 132 of flow rate indicator 25 will be in its fully raisedposition as shown in FIG. 2B.

In the preferred form of control linkage shown in FIGS. 2A-C, theelectrical contact 136 is mounted on a link 167 which is connected bylinks 168 and 169 to the accelerator control pedal 151. The function oflinks 167, 168 and 169 is to adjust the position of electrical contact136 so that conversion from four-wheel drive to two-wheel drivewilloccur at a vehicle speed which is approximately equal to half the finalspeed which will be reached for a particular position of the acceleratorcontrol pedal 151. For example, if the accelerator pedal 151 is fullydepressed, the final speed of the vehicle will be its maximum speed, andconversion from four-wheel drive to two-wheel drive will occur atapproximately half maximum speed. On the other hand, if the acceleratorpedal 151 is depressed only halfway, the final speed of the vehicle willbe approximately half of its maximum speed, and conversion fromfour-wheel drive to two-wheel drive will occur at approximatelyone-quarter of maximum speed.

It will be appreciated that, as a vehicle equipped with the presenthydraulic power transmission system decelerates from a cruising speed,low rate indicator 25 will operate to automatically convert the vehiclefrom twowheel drive back to four-wheel drive at a speed which isdetermined by the position of the accelerator control pedal 151.

Referring now to FIG. 2C of the drawings, there is shown a diagram ofthe control apparatus in the RE- VERSE condition. The movable member 132of flow rate indicator 25 is in its lowest position, corresponding to azero flow or a reverse flow of hydraulic fluid through passage 131, sothat stop 171 mounted on movable member 132 contacts L-shaped link 172which pivots about a fixed pin 173 and is connected at its opposite endto a link 174 which is connected by a tension spring 175 to link 176which is connected to the selector level 101. When movable member 132 isin its lowest position, as shown in FIG. 2C, stop 171 causes L-shapedlink 172 to be rotated in the counterclockwise direction about pin 173,thereby drawing links 174 and 176 to the left against the action oftension spring 177, thereby causing the lower end 176a of link 176 toclear fixed pin 178, so that, when selector 101 is moved to the REVERSEposition as shown in FIG. 2C, the end 176a of link 176 bears against Lshaped link 181 and causes it to pivot in the clockwise direction aboutthe fixed pin 182 against the action of tension spring 184. The motionof L-shaped link 181 is transmitted by link 183 to link 156 thus causingpin 155 which is mounted on link 156 to move to the left end 185a ofslot 185 in accelerator control pedal 151 as shown in FIG. 2C.

It will be appreciated that, if the movable member 132 of flow rateindicator 25 is not in its lowest position as shown in FIG. 2C, selectorlever 101 will not be able to move to the REVERSE position, because end176a of link 176 will bear against the stationary pin 178 rather thanL-shaped link 181. Hence, the effect of stop 171 on L-shaped link 172insures that the present hydraulic transmission system cannot be putinto the RE- VERSE condition while the vehicle is moving forward.

When the control apparatus of the present hydraulic power transmissionsystem is in the REVERSE condition, as shown in FIG. 2C, and theaccelerator control pedal 151 is depressed, link 156 moves upward, thuscausing link 157 to pivot in a clockwise direction about pin 161. Theresulting upward motion of the opposite end of link 157 is transmittedthrough links 152 and 153 and the pilot valve member 108 of main c0ntrolvalve 22. When the pilot valve member 108 is moved to its raisedposition, as shown in FIG. 2C, hydraulic fluid from the low pressure(150 p.s.i.) system flows through passages 111 and 113 to the upperportion 115 of the main valve chamber, thus causing the main valvemember 109 to move downward. It will be appreciated that the downwardmovement of main valve member 109 is coupled through link 163 to pilotvalve member 108, thus tending to centralize the position of pilot valve103.

When the main control valve member 109 is in its lower (or REVERSE)position as shown in FIG. 2C, high pressure hydraulic fluid from thevariable dis placement pump flows through conduit 20, passage 125 andconduit 24 to the hydraulic wheel motors, while hydraulic fluidreturning from the wheel motors via line 23 flows through passage 123and conduit 44 back to the intake side of the variable displacementhydraulic pump 21. Thus it is seen that the operation of the maincontrol valve 22 serves to direct hydraulic fluid to the wheel motorsvia hydraulic line 24, with the return flow being carried by hydraulicline 23 when the selector lever 101 is in the REVERSE position as shownin FIG. 2C, while hydraulic fluid is directed to the wheel motors vialine 23 with the return flow being carried by line 24 when the selectorlever 101 is in the FORWARD position, as shown in FIG. 28.

Although the control apparatus illustrated in FIGS. 2A-C uses thehydraulic fluid flow rate indicator 25 to effect switching betweenfour-wheel drive and twowheel drive, it will be appreciated that othertypes of devices may be used to perform this function. For example,switching between four-wheel and two-wheel drive might be accomplishedby a conventional mechanical governor device coupled to the wheels ofthe vehicle.

The Hydraulic Wheel Motors Referring to FIG. 3 of the drawings, there isshown a cross-sectional view, taken in a plane perpendicular to the axisof rotation of one of the fixed displacement hydraulic wheel motors usedin the preferred form of the present hydraulic power transmission andbraking system. The hydraulic motor of FIG. 3 is of the type in whichthe driven element is in the form of an orbiting spider, generallydesignated 201, which is eccentrically mounted on a driven shaft 202.The legs 204, 205, 206, 207 and 208 of the spider 201 perform severalfunctions including (a) covering and uncovering the inlet ports 211,212, 213, 214 and 215 and the outlet ports 221, 222, 223, 224 and 225 inthe proper sequence, and (b) preventing the spider 201 from rotating andthus constraining it to orbital motion.

The hydraulic wheel motor of the FIG. 3 includes a cylindrical casing230 which includes castings 238 and 239 and circumferential member 240.Cylindrical casing 230 has five circumferentially spaced chambers 231,232, 233, 234 and 235 which change in size as the spider 201 orbitsabout the shaft 202. Each of the chambers 231-235 is bounded by aportion of the spider 201, portions of the opposed end surfaces 236 and237 of castings 238 and 239 (shown more clearly in FIG. 4), a portion ofthe circumferential member 240 and two of the movable vanes 241, 242,243, 244 and 245. The vanes 241-245 form seals between the chambers231-235 of the hydraulic motor shown in FIG. 3. For example, chamber 231is separated from chamber 235 by vane 241, and separated from chamber232 by vane 242.

In the preferred form of hydraulic motor shown in FIG. 3, the inner endsof vanes 241-245 are mounted to the spider 201 by a cylinder and socketarrangement which permits the vane to pivot from side to side withrespect to the spider 201 while maintaining a reasonable good fluid sealbetween the vane and the spider. The outer ends of the vanes 241-245 aremovably mounted to the circumferential casing 240 by means of splitcylinders 251, 252, 253, 254 and 255 respectively. Small cavities 261,262, 263, 264 and 265 are provided in circumferential casing 240 toallow free movement of the outer ends of vanes 241-245 as the spider 201orbits around shaft 202. Passages 271, 272, 273, 274 and 275 connectcavities 261-265 to chambers 231-235 to relieve the pumping pressurecaused by the movement of the outer ends of vanes 241-245 within thecavities 261-265 as the spider 201 orbits the shaft 202.

Referring now to FIG. 4 of the drawings which shows a cross-sectionalview of the preferred form of hydraulic wheel motor taken along the line4-4 of FIG. 3, it will be seen that all of the inlet ports 211-215 ofFIG. 3 are connected together by a manifold arrangement within casting238. More specifically, all of the inlet ports 211-215 are connected toa circumferential passage 276 within casting 238. Similarly, all of theoutlet ports 221-225 of FIG. 3 are connected to the circumferentialpassage 277 within casting 238. Similarly, the inlet openings in surface237 of casting 239 are connected by the circumferential passage 278, andthe outlet openings in surface 237 of casting 239 are connected by thecircumferential casting 279. It will be appreciated that the inletopenings in surface 237 of casting 239 are located opposite the inletopenings 211-215 in the surface 236 of casting 238, and the outletopenings in the surface 237 of casting 239 are located opposite theoutlet openings 221-225 in the surface 236 of casting 238.

Referring to FIG. 5 which is a cross-sectional view taken along the line5-5 in FIG. 3, it will be seen that the circumferential outlet passage277 in casting 238 is connected to the circumferential outlet passage279 in casting 239 by a suitable Y coupling 281 while thecircumferential inlet passage 276 :in casting 238 is connected to thecircumferential inlet passage 278 in casting 239 by a similar Y coupling282 (shown in FIG. 3) which is connected to the coupling point 202A and282B shown in FIG. 5. Thus, Y coupling 281 serves to connect all theoutlet ports in the hydraulic wheel motor to a single hydraulic linewhile the Y coupling 282 serves to connect all the inlet ports in thehydraulic wheel motor to a single hydraulic line.

Assuming that the hydraulic wheel motor will drive the vehicle in theforward direction when the high pressure hydraulic line from thevariable displacement pump 21 (FIG. 1) is connected to the inlet portsof the hydraulic motor of FIGS. 3-5 via the Y coupling 282 and thereturn hydraulic line to the variable displacement pump 21 is connectedto the outlet ports of the hydraulic motor via the Y coupling 281, itwill be appreciated that the motor of FIGS. 3-5 can be made to drive thevehicle in the reverse direction simply by connecting the high pressureline from the pump 21 to Y coupling 281 and connecting the return lineto the pump 21 to the Y coupling 282. This reversing action isaccomplished by the main. control valve 22 described in connection withFIGS. 1 and ZA-C.

Referring again to FIG. 4- of the drawings, it will be seen that, in thepreferred embodiment, the castings 238 and 239 and the circumferentialmember 240 are held together by means of suitable fastening devices suchas, for example, bolts 283. O-ring seals 284 and 285 are providedbetween circumferential member 240 and castings 238 and 239 in order toprevent leakage of hydraulic fluid. An end plate 286 is fastened tocasting 238 by means of suitable fastening devices such as bolts 287,and an O-ring seal 288 is provided between end plate 286 and casting 238to prevent hydraulic fluid leakage. A ring gear carrier 289 is mountedon casting 239 by means of suitable fastening devices such as forexample bolts 290, and again an O-ring seal 291 is provided to preventhydraulic fluid leakage.

Mechanical torque is transmitted from the orbitting spider 201 throughthe eccentric 203 to the shaft 202 which is rotatably mounted withincastings 238 and 239 by means of roller bearings 293 and 294respectively. One end of shaft 202 is rotatably mounted to end plate 286by a thrust bearing 295 while the other end of shaft 202 carries a sungear 296 which engages planetary gears 297 which are rotatably mountedon a planetary gear carrier 298 which is rotatably mounted with ringgear carrier 289 by a suitable ball bearing 299. A pair ofcounter-weights 301 and 302 are splined to shaft 202 to counter balancethe eccentric mass of the spider 201. The vehicle wheel is connected tothe end of shaft 303 projecting from the planetary gear carrier 298.Port 304 permits hydraulic fluid to be scavenged from the spaces withinthe hydraulic motor of FIGS. 3-5.

Briefly, the operation of the preferred form of the hy draulic wheelmotor shown in FIGS. 3-5 is as follows. Hydraulic fluid under pressureenters chambers 231 and 232 through inlet ports 211 and 212respectively. The pressure in chambers 231 and 232 tends to drive thespider 201 in the clockwise direction thus driving the shaft 202 througheccentric 203. As the spider 201 moves in the clockwise direction, thevolume of chambers 234 and 235 is reduced thus causing hydraulic fluidto flow out through outlet ports 224 and 225 respectively. At the sametime, leg 206 of spider 201 begins to uncover inlet port 213 of chamber233 while leg 204 of spider 201 begins to cover both inlet port 211 andoutlet port 221 of chamber 231.

When the inlet port 211 of chamber 231 is closed by leg 204 and theinlet port 213 of chamber 233 is open, the high pressure in chambers 232and 233 will continue to move the spider 201 in the clockwise direction.It will be apparent that further clockwise orbital motion of spider 201will continue to cover and uncover inlet and outlet ports in the propersequence to bring about the continued clockwise orbiting of the spider201 with the motion of the spider 201 being transmitted to the shaft 202through the eccentric 203. Referring now to FIG. 4 of the drawings, itwill be seen that the rotation of shaft 202 causes the sun gear 296 todrive the planetary gears 297 which are mounted on the planetary gearcarrier 298. In the preferred embodiment of the present invention, theplanetary gear arrangement accomplishes a 3:1 speed reduction betweenthe shaft 202 and the shaft 303 which is connected to the vehicle wheel.

The Variable Displacement Hydraulic Pump Referring now to FIG. 6 of thedrawings, there is shown a cross-sectional view, taken in a planeperpendicular to the axis of rotation, of the preferred form of variabledisplacement hydraulic pump 21 used in the hydraulic power transmissionand braking system of the present invention. Comparing the hydraulicpump 21 of FIG. 6 with the hydraulic motor of FIG. 3, it will beapparent that there are certain similarities of structure. For example,the variable displacement hydraulic pump of FIG. 6 includes a spider 308mounted for clockwise eccentric orbital'movement with a cylindricalcasing 307. The spider 308 has five legs 311-315 which serve to coverand uncover inlet ports 321-325 and outlet ports 331-335 in the propersequence. The pumping chambers 341-345 are separated by movable vanes351-355.

It will be appreciated, however, that the variable displacementhydraulic pump 21 of FIG. 6 differs in certain important respects fromthe fixed displacement hydraulic motor of FIG. 3. More particularly, theapparatus of FIG. 6 operates as a pump rather than as a motor. That is,the spider 308 is driven by a suitable prime mover, such as for examplean internal combustion engine, to cause hydraulic fluid to be drawn inthrough inlet ports 321-325 and expelled through outlet ports 331-335.Second, the pump 21 of FIG. 6 is a variable displacement device, thedisplacement of which is controlled by the degree of eccentricity of theorbital movement of spider 308 within the casing 307. The displacementof the hydraulic pump of FIG. 6 is greatest when the degree ofeccentricity of spider 308 is greatest, and the displacement of the pumpis least when the degree of eccentricity of spider 308 is least. It willbe appreciated by those skilled in the art that, for a constant powerinput from the prime mover, the output pressure of the variabledisplacement hydraulic pump 21 of FIG. 6 will be inversely proportionalto its effective displacement.

Referring now to FIG. 7 of the drawings, there is shown across-sectional view of the variable displacement hydraulic pump 21taken along the line 7-7 of FIG. 6. The spider 308 is located within thespace bounded by circumferential member 309 and the opposed surfaces 361and 362 of castings 363 and 364 respectively. Circumferential member 309and castings 363 and 364 are held together by suitable fastening meanssuch as, for example, bolts 365 to form the cylindrical casing 307.O-rings 366 and 367 are provided between casing 309 and castings 363 and364 respectively in order to prevent leakage of hydraulic fluid. Thelegs of spider 308, such as for example leg 313 shown in FIG. 7, arepreferably hollow in order to reduce the eccentric mass of the spiderand thus reduce the dynamic balance problems of the pump.

As in the case of the hydraulic wheel motor shown in FIGS. 3-5 all ofthe inlet ports in surface 361 of casting 363, such as inlet port 323for example, are connected to a circumferential passage 371 in casing363 while all of the inlet ports in surface 362 of casting 364 areconnected to the circumferential passage 372 in casting 364.Circumferential passages 371 and 372 are connected together by a Ycoupling 373, shown in FIG. 6, and Y coupling 373 is connected to thereturn hydraulic line from the wheel motors.

Similarly, all of the outlet openings in surface 361 of casting 363 areconnected to the circumferential passage 375 while all of the outletopenings in surface 362 of casting 364 are connected to thecircumferential passage 376. The two circumferential passages 375 and376 are connected together by a suitable Y coupling 377, shown in FIG.6, and Y coupling 377 is connected to the output hydraulic line to thewheel motors.

Briefly, power is transmitted from the prime mover through an outereccentric carrier 381, and inner eccentric carrier 382, and eccentricshaft 383 and the spider center member 384 to the spider 308. The spider

1. A hydraulic power transmission system for vehicles comprising: ahydraulic pump having an outlet port and an inlet port for hydraulicfluid; a plurality of hydraulic motors, each hydraulic motor having afirst port and a second port for hydraulic fluid and a shaft connectedto a wheel of the vehicle; a control valve connected between saidhydraulic pump and said hydraulic motors, said control valve having afirst position connecting said outlet port of said hydraulic pump tosaid first ports of said hydraulic motors and connecting said inlet portof said hydraulic pump to said second ports of said hydraulic motors,and a second position connecting said outlet port of said hydraulic pumpto said second ports of said hydraulic motors and said inlet port ofsaid hydraulic pump to said first ports of said hydraulic motors; apilot valve for controlling the movement of said second control valvebetween said first position and said second position, said pilot valvebeing connected in a follow-up relation with said control valve; aselector member controllable by a vehicle operator, said selector memberhaving a FORWARD position and a REVERSE position; an accelerator membercontrollable by a vehicle operator for controlling the speed of saidhydraulic pump; and control means connecting said selector member andsaid accelerator member to said pilot valve and said control valve formoving said control valve toward said first position when said selectormember is in said FORWARD position and said accelerator member is movedto increase the speed of said hydraulic pump, and for moving saidcontrol valve toward said second position when said selector member isin said REVERSE position and said accelerator member is moved toincrease the speed of said hydraulic pump.
 2. The hydraulic powertransmission system of claim 1 wherein said control valve has a thirdposition connecting said outlet port of said hydraulic pump to saidinlet port of said hydraulic pump and connecting said first ports ofsaid hydraulic motors to said second ports of said hydraulic motors; andwherein said selector member has a NEUTRAL position; and wherein saidcontrol means moves said control valve to said third position when saidselector member is in said NEUTRAL position.
 3. A hydraulic powertransmission system for vehicles comprising: a variable displacementhydraulic pump having an outlet port and an inlet port for hydraulicfluid; a plurality of hydraulic motors, each hydraulic motor having afirst port and a second port for hydraulic fluid and a shaft connectedto a wheel of the vehicle; a control valve connected between saidhydraulic pump and said hydraulic motors, said control valve having afirst position connecting said outlet port of said hydraulic pump tosaid first ports of said hydraulic motors and connecting said inlet portof said hydraulic pump to said second ports of said hydraulic motors,and a second position connecting said outlet port of said hydraulic pumpto said second ports of said hydraulic motors and said inlet port ofsaid hydraulic pump to said first ports of said hydraulic motors; aselector member controllable by a vehicle operator, said selector memberhaving a FORWARD position and a REVERSE position; an accelerator membercontrollable by a vehicle operator for controlling the speed of saidhydraulic pump; control means connecting said selector member and saidaccelerator member to said control valve for moving said control valvetoward said first position when said selector member is in said FORWARDposition and said accelerator member is moved to increase the speed ofsaid hydraulic pump, and for moving said control valve toward saidsecond position when said selector member is in said REVERSE positionand said accelerator member is moved to increase the speed of saidhydraulic pump; and displacement control means for increasing thedisplacement of said variable displacement hydraulic pump in response toincreasing speed of said hydraulic wheel motors and for decreasing thedisplacement of said variable displacement hydraulic pump in response todecreasing speed of said hydraulic wheel motors.
 4. The hydraulic powertransmission system of claim 3 wherein said displacement control meansoperates to increase the displacement of said variable displacementhydraulic pump in response to decreasing torque on said pump and todecrease the dsiplacement of said variable displacement hydraulic pumpin response to increasing torque on said pump.
 5. The hydraulic powertransmission system of claim 4 wherein said displacEment control meansoperates to increase the displacement of said variable displacementhydraulic pump in response to increasing speed of said pump and todecrease the displacement of said variable displacement hydraulic pumpin response to decreasing speed of said pump.
 6. The hydraulic powertransmission system of claim 3 further comprising: a recirculating valveconnected between said control valve and at least one of said hydraulicmotors, said recirculating valve having a first position connecting theflow of hydraulic fluid from said control valve through said hydraulicmotor and back to said control valve, and a second position cutting offthe flow of hydraulic fluid from and toward said control valve andconnecting said first port of said hydraulic motor to said second portof said hydraulic motor to permit hydraulic fluid to circulate freelythrough said hydraulic motor.
 7. The hydraulic power transmission systemof claim 6 further comprising: means responsive to the speed of thevehicle for moving said recirculating valve to said first position whenthe vehicle speed is below a predetermined level and for moving saidrecirculating valve to said second position when said vehicle speedexceeds said predetermined level.
 8. A hydraulic power transmissionsystem for vehicles comprising: a hydraulic pump having an outlet portand an inlet port for hydraulic fluid; a plurality of hydraulic motors,each hydraulic motor having a first port and a second port for hydraulicfluid and a shaft connected to a wheel of the vehicle; a control valveconnected between said hydraulic pump and said hydraulic motors, saidcontrol valve having a first position connecting said outlet port ofsaid hydraulic pump to said first ports of said hydraulic motors andconnecting said inlet port of said hydraulic pump to said second portsof said hydraulic motors, and a second position connecting said outletport of said hydraulic pump to said second ports of said hydrauliicmotors and said inlet port of said hydraulic pump to said first ports ofsaid hydraulic motors; a selector member controllable by a vehicleoperator, said selector member having a FORWARD position and a REVERSEposition; an accelerator member controllable by a vehicle operator forcontrolling the speed of said hydraulic pump; control means connectingsaid selector member and said accelerator member to said control valvefor moving said control valve toward said first position when saidselector member is in said FORWARD position and said accelerator memberis moved to increase the speed of said hydraulic pump, and for movingsaid control valve toward said second position when said selector memberis in said REVERSE position and said accelerator member is moved toincrease the speed of said hydraulic pump; a recirculating valveconnected between said control valve and at least one of said hydraulicmotors, said recirculating valve having a first position connecting theflow of hydraulic fluid from said control valve through said hydraulicmotor and back to said control valve, and a second position cutting offthe flow of hydraulic fluid from and toward said control valve andconnecting said first port of said hydraulic motor to the second port ofsaid hydraulic motor to permit hydraulic fluid to circulate freelythrough said hydraulic motor, and means for indicating the rate of flowof hydraulic fluid through at least one of said hydraulic motors; andactuating means responsive to said indicating means for moving saidrecirculating valve to said first position when said flow rate is belowa predetermined level, and for moving said recirculating valve to saidsecond position when said flow rate exceeds said predetermined level. 9.The hydraulic power transmission system of claim 8 wherein sad actuatingmeans further comprises: means for adjusting said predetermined flowrate level in response to the position of said accelerator member. 10.The hydraulic power trAnsmission system of claim 8 further comprising:means responsive to said flow rate indicating means for inhibiting saidselector member from being moved to the REVERSE position unless thehydraulic fluid flow rate is substantially zero.
 11. The hydraulic powertransmission system of claim 8 further comprising: means responsive tosaid flow rate indicating means for increasing the actuation of saidcontrol valve in response to increasing hydraulic fluid flow rate. 12.The hydraulic power transmission system of claim 1 further comprising apower source for driving said hydraulic pump, wherein said acceleratormember controls the speed of said power source.
 13. The hydraulic powertransmission system of claim 1 further comprising: a forward brake valveconnected between said inlet port of said hydraulic pump and the secondport of at least one of said hydraulic motors; and a brake pedal foractuating said forward brake valve to progressively restrict the returnflow of hydraulic fluid from said hydraulic motor to said hydraulicpump, thus providing a braking effect to said hydraulic motor.
 14. Thehydraulic power transmission system of claim 13 further comprising: areverse brake valve connected between said inlet port of said hydraulicpump and the first port of at least one of said hydraulic motors; fluidflow direction sensing means for sensing hydraulic fluid flowcorresponding to movement of the vehicle in the reverse direction; andmeans responsive to said flow direction sensing means for causing saidbrake pedal to actuate said forward brake valve when the fluid flowdirection corresponds to forward vehicle movement and for causing saidbrake pedal to actuate said reverse brake valve when said fluid flowdirection corresponds to reverse vehicle movement.
 15. The hydraulicpower transmission system of claim 1 further comprising: a plurality offorward brake valves, each forward brake valve being connected betweenthe second port of one of said hydraulic motors and said control valve;a plurality of reverse brake valves, each reverse brake valve beingconnected between the first port of one of said hydraulic motors andsaid control valve; fluid flow direction sensing means; a brake pedalcontrollable by a vehicle operator; and brake control apparatus foractuating said forward brake valves in response to said brake pedal whensaid fluid flow direction sensing means senses fluid flow correspondingto forward vehicle movement, and for actuating said reverse brake valvesin response to said brake pedal when said fluid flow sensing meanssenses fluid flow corresponding to reverse vehicle movement.
 16. Thehydraulic power transmission system of claim 15 wherein said brakecontrol apparatus comprises: a plurality of forward brake controlcylinders connected to said forward brake valves; a plurality of reversebrake control cylinders connected to said reverse brake valves; meanslinking said brake pedal to each of said brake control valves, saidlinking means being responsive to said fluid flow direction sensingmeans to cause said brake pedal to actuate said forward brake controlcylinders in response to said brake pedal when said fluid flow directionsensing means senses fluid flow corresponding to forward vehiclemovement, and for actuating said reverse brake control cylinders inresponse to said brake pedal when said fluid flow sensing means sensesfluid flow corresponding to reverse vehicle movement.
 17. The hydraulicpower transmission system of claim 16 further comprising: a pendulumconnected to said linking means for increasing the actuation of one ofsaid brake control cylinders and decreasing the actuation of another ofsaid brake control cylinders when said pendulum is displaced from itsnormal position.
 18. The hydraulic power transmission system of claim 15further comprising: an overpressure relief valve connected in parallelwith each of said hydraulic mOtors for permitting hydraulic fluid toflow between the first port and the second port of said hydraulic motorwhen the differential hydraulic fluid pressure between said first andsecond ports exceeds a predetermined value.
 19. The hydraulic powertransmission system of claim 15 further comprising: means for scavenginghydraulic fluid from said hydraulic motors and said hydraulic pump; anauxiliary hydraulic pump connected to said scavenging means for pumpingthe scavenged hydraulic fluid to a predetermined pressure; and at leastone check valve connecting said auxiliary hydraulic pump to a point in ahydraulic line between said hydraulic pump and said hydraulic motors forsupplying hydraulic fluid to said hydraulic line when the pressure insaid hydraulic line falls below said predetermined pressure butpreventing hydraulic fluid from flowing from said hydraulic line to saidauxiliary hydraulic pump.
 20. A hydraulic power transmission and brakingsystem for vehicles comprising: a hydraulic pump having an outlet portand an inlet port for hydraulic fluid; a plurality of hydraulic motors,each hydraulic motor having a first port and a second port for hydraulicfluid and a shaft connected to a wheel of the vehicle; a control valveconnected between said hydraulic pump and said hydraulic motors forreversibly controlling the flow of hydraulic fluid from said hydraulicpump to said hydraulic motors; a brake pedal controllable by a vehicleoperator; a plurality of forward brake valves connected between saidsecond ports of said hydraulic motors and said control valve forrestricting the return flow of hydraulic fluid from said hydraulic motorto said hydraulic pump when the vehicle is moving in the forwarddirection; a plurality of reverse brake valves connected between saidfirst ports of said hydraulic motors and said control valve forrestricting the return flow of hydraulic fluid from said hydraulicmotors to said control valve when the vehicle is moving in the reversedirection; means for sensing the direction of hydraulic fluid flowbetween said hydraulic pump and said hydraulic motors; means linkingsaid brake pedal to each of said brake valves for actuating said brakevalves in response to said brake pedal, said linking means includingmeans responsive to said fluid flow direction sensing means for causingsaid brake pedal to actuate only said forward brake valves when thefluid flow direction corresponds to forward vehicle movement and forcausing said brake pedal to actuate only said reverse brake valves whenthe fluid flow direction corresponds to reverse vehicle motion; and apendulum connected to said linking means for increasing the actuation ofat least one of said brake valves and decreasing the actuation of atleast one other of said brake valves when said pendulum is displacedfrom its normal position.